DE69922504T2 - MOTOR VEHICLE GEARBOX - Google Patents

Abstract

A power dividing device for a vehicle transmission to provide two rotating outputs having variable relative speed of rotation from a single rotatable input. The device includes a first rotational element and a second rotational element housed within the first element and rotatable about the same axis as the second rotational output. A first fluid chamber associated with the first rotational element and first regulating means to varies the volume of the first chamber in response to rotation of the first rotational element. A second fluid chamber associated with the second rotational element and second regulating means likewise varies the volume to the flow between the first and second chambers the relative timing of variation of the volumes of the first and second chambers determines the speed of the second rotational element in response to rotation of the first rotational element.

Description

AREA OF INVENTION

The The invention relates to motor vehicle drive trains.

STATE OF TECHNOLOGY

Automotive powertrains exist generally from a source of energy via a coupling mechanism connected to a transmission, manually or by means of a Series of clutches and brakes is controlled, with in both make in general, three or more different drive ratios to disposal stand. Most motor vehicles still use an internal combustion engine as an energy source. In general, internal combustion engines have a narrow Operating speed bandwidth (RPM) based on the torque or the driving force is tuned as a preferred operating point at maximum fuel efficiency and minimal pollution results.

The Effect of the stepped transmission is that each of the gears or Gradations of the engine first exposed to a high torque requirement at low speeds is, causing a stall is caused when the piston moves slower than the flame front in the combustion stroke. When the speed of the engine increases, happens he the preferred operating point of balanced operation of Speed and torque, with the piston moving at speed the flame front and the expanding air moves. Then exceeds the engine speed is the balance between speed and torque demand and gets into an overspeed situation, until the next one Gear is engaged and the above is repeated during the next and subsequent gear shifts becomes.

On open country road let yourself greater efficiency through an overdrive gear reach the speed of the engine and the torque request again better coordinated. Usually it is thereby to a fixed transmission ratio and in general, almost the maximum performance of the engine for operation on even or moderately uneven ground Terrain. at a slope must the Driver with most automatic transmissions in a relatively large gradation the resulting shortcomings one Manually press the button or automatically downshift to a lower gear ratio. If the load is in the middle of the highest gear / overdrive, Switches the gearbox in regular, relative fast cycles frequently several times up and down, resulting in one for both the driver as well as for the passengers noticeable Acceleration and deceleration results.

At the Attempt to keep the motor constant and steady at the preferred operating point the state of compensation of engine speed and required torque and that with the combustion cycle flame front and expanding Gas-balanced piston speed to operate, shows the PRIOR ART Attempts to Combine Various Transmission Systems which require a variable speed drive. In general use conical friction drive pulleys or pulleys, where the speed change by sliding the tapered cones relative to each other is reached. The amount of torque transferable in this way is minimal and the consequent wear makes this option unusable. The success of V-belt pulleys and matching adjustable V-belt pulleys using a wide V-belt were low in the low horsepower range.

hydraulic pumps and variable speed variable speed gearboxes in use The technology is designed so that with increasing speed the output also increases the amount of oil flow, so that due whose further loss of friction is made. When trying Minimizing this loss has been both higher pressures and lower flow rates used. This has very poor torque characteristics in the low-speed Area result. The resulting loss of efficiency and performance along with energy loss, heat generation and necessity the cooling through additional circuits and cooler proves that this Gearbox form for the use in high-speed vehicles is not useful and only used in low-speed tractors and mobile machines.

Applicant's International Patent Application WO 98/17972 discloses a vehicle load transmission and drive train capable of being continuously controlled over a predetermined range of effectiveness by means of an external transmission and an internal differential gear assembly controlled by two rotatable drives. The inputs from a microprocessor continuously monitor all critical inputs and outputs and provide continuous real-time micro-adjustments to guarantee the driver and passengers a comfortable and quiet ride, fuel efficiency and pollution reduction. Although a single energy source, eg hydraulic drive systems, can be used to drive the two drives, the International Application mainly describes the use of a part engine or two energy sources, wherein the one engine for maximum fuel and pollution benefits continuously runs at the preferred operating point, while the second power source is used to compensate for the variable driving forces. The disclosure of International Application WO 98/17927 hereby forms part of this application by cross-reference and will be referred to hereinafter as "the Applicant's earlier application".

GB 1161508 describes a variable transmission system according to the preamble of claim 1 having a single drive shaft connected to a sun gear and a planetary gear 8th connected is. The sun wheel 8th drives a wave 11 on, with a pump 3 connected is. The pump 3 transfers energy to an engine 4 that drives the output shaft 37 drives. The planet wheel 10 transmits via a clutch 46 and gears 38 . 39 Energy to a decline wave 42 , However, this system requires a clutch system to provide a variable mechanical transmission. In addition, the system can not be adapted to existing vehicles.

US 3,698,189 describes a neutral hydraulic clutch valve for a hydromechanical transmission. The valve is used in a hydromechanical unit having only a single input connected to a pump unit and a single output connected to a motor unit. The housing of the hydromechanical unit is integrally formed with the outlet and rotatable together with the pump unit. The hydromechanical unit of this patent does not allow splitting the power input to provide a variable speed output, and thus does not maximize engine efficiency.

EPIPHANY THE INVENTION

For mass production of automobiles and to minimize tooling investment costs the use of existing mass-production machines is an advantage. It is also desirable a complete singular Energy source continuously with maximum balanced speed and operate the same torque, and constant in the preferred operating point for maximum fuel efficiency and minimal pollution, and these Operating in all spectrums of the range of driving a vehicle maintain.

A It is an object of one aspect of this invention to provide a transfer case device. two rotary outputs with a variable relative speed from a single rotary drive. A task of a Another aspect of the present invention is a Vehicle powertrain and a powershift transmission to provide over a predetermined operating range can be controlled continuously.

Under In a first aspect, the present invention provides a transfer case device Provision of two rotary drives with a variable relative Speed from a single rotary drive, wherein the device comprising: a first rotary element rotating in rotation a rotation axis is driven and a first rotation output having; a second rotation member that is rotatable about the rotation axis is and has a second rotational output; a first fluid chamber, which belongs to the first rotation element; a first controller for Varying the volume of the first chamber in response to the rotation the first rotary element; a second fluid chamber leading to the heard second rotary element; a second regulator for varying the volume of the second chamber in response to the rotation of the second rotary member; thereby characterized in that a commutator in response to the rotation of the first rotation element is rotatable to at least one regular closed fluid flow communication between the first and the second chamber during to produce the rotation of the first and second rotation elements; wherein the rotational speed of the second rotational element in response to the rotation of the first rotation element by corresponding changes the volume of the first and second chamber is determined.

In another aspect, this invention provides a power transmission unit comprising: a transfer case, an outer main gear, and an inner differential gear assembly;
the main gearbox having two rotary drive means each driven by a first rotary output and a second rotary output gear of the transfer gear device, the two drive means being operatively connected to the rotary output means such that the rotational speed of the output means is proportional to the algebraic average of the rotational speed both drive means may vary;
the differential gear assembly being disposed in the main gear and having rotatable drive means operatively connected to two differentially rotatable output means, the output gear of the main gear and the drive means of the differential gear assembly being operatively connected; characterized in that the transfer case device according to the first aspect is arranged.

In another aspect, the present invention provides an automotive powertrain capable of being continuously controlled over a predetermined operating range and comprising:
a single engine;
a power transmission unit comprising a transfer gear device, an outer main gear, and an inner differential gear assembly;
the main gearbox having two rotatable drive means each driven by a first rotational output and a second rotational output gear of the transfer gear device, the two drive means being operatively connected to the rotatable output means such that the rotational speed of the output means is proportional to the algebraic average of the rotational speed the two drive means can vary;
the differential gear assembly being disposed in the main gearbox and having rotatable drive means operatively connected to two differentially rotatable output means, the output gear of the main gear and the drive means of the differential gear assembly being operatively connected; characterized in that the transfer case device according to the first aspect is arranged.

Supplies according to the invention a single engine over a drive train a drive branch at a constant speed and over a second drive train a drive branch of the same engine, whereby it is possible by means of closed fluid volumes or columns, preferably oil, that the reaction forces of Gear directly without energy loss when balancing the forces the energy source react.

Preferably, the second drive branch mechanically varies the amounts of trapped oil to allow the speed of the two drives for the transmission to vary, thereby reducing the output speed according to the formula Vout = 2 × Vsecondary - Vprimary Vout is the output speed, Vsecondary is the speed of the secondary drive provided by the second rotary output of the transfer case, and Vprimary is the speed of the prime mover provided by the first rotary output of the transfer case.

According to the invention, the transfer case utilizes a unique method of proportional control of the trapped oil such that the second rotational output is preferably always:
in the same direction as the first rotation output;
never less than one third of the rotational speed of the first rotary output when the transmission output is in the reverse gear unless it is used on machines that require the same speed in forward and reverse, in which case the rotational output is zero U / min can be reduced;
at half the speed of the first rotary output rotates when in neutral dynamic lock;
running at the same speed as the first rotation output when in full forward motion.

The differential gear assembly is internally disposed in the main gearbox and has rotational drive means operatively connected to two differential rotational output members.
wherein the output of the main gear and the drive means of the differential gear assembly are operatively connected to each other.

It Control devices are provided which have a device for receiving a command input and means for determining the with powertrain operation related performance parameters exhibit. The performance parameters include the load of the engine, wherein the pressure of the encapsulated oil columns in the secondary drive branch the forces of the external gear limited, the speeds of the first primary drive branch and the secondary drive branch of the Main gear, the respective load of the two output means the differential gear assembly and the respective speeds the two output means of the differential gear assembly.

The Control device ensures a feedback loop control for the continuous monitoring, Analysis and adjustment of performance parameters in response to the Command input.

The Engine can be made from any conventional internal combustion engine exist - including the Otto engine, the diesel engine, a rotary engine including one true balanced rotary piston engine with compression and expansion cycle or a gas turbine engine as well as conventional types of electric motors. The Otto and diesel internal combustion engines are preferred because they established mass production technologies with relatively low production costs represent.

Preferably rotate the Rotationsleistungsabtriebsmittel the primary and secondary drive branches with respect to the power transmission unit in the same direction.

Preferably, at least one of the rotational power take-off means is the two power operating branches selectively connected via a coupling means operatively connected to the drive means of the main transmission of the power transmission unit. Alternatively, the microprocessor, which detects zero input signals at idle, issues commands for positioning the encapsulated oil column control mechanism to set the secondary drive branch speed to one-half that of the transmission prime gear input so that a 2: 1 drive ratio is zero output rotation generated with dynamic locking in the zero output position.

The effective connection can also be gear, chain, belt, electric, hydraulic or direct motor drive shaft connection devices.

The Engines, the primary and the secondary drive, the outer gear and the inner differential can suitably surrounded by a common housing or expediently be detachable as an assembly of the engine.

Of the Drive train of this invention allows the optimization of Engine design for different applications. An electric motor with constant speed For example, as a generator for regenerative braking and Slowing down to be used. An engine of any type, that is at constant maximum efficiency, either directly or over a high speed flywheel enclosed in a vacuum is and can run on low-friction bearings or magnetic bearings with highly variable power requirements and regenerative braking be used, for example, on city trips, so that a Hybrid with extremely low pollution results.

The main transmission of the powershift transmission unit advantageously has a gear train. Preferably, the two drive means of the main gear on a first bevel gear and a planetary gear arrangement, which are arranged rotatably coaxially about a first axis. Advantageously, the epicyclic gear arrangement has an annular pinion carrier which rotatably supports the internally disposed epicyclic pinions whose axes are perpendicular to the first axis. Conveniently, the first bevel gear and the annular pinion carrier are each operatively individually with rotating energy sources, such as the primary and secondary drive branches of a singular energy source or a combination of a hybrid power source with a singular output shaft, the transmission unit both on the primary and on the secondary drive branch supplies energy, operatively connected. Preferably, the output means of the main gear comprises a second bevel gear coaxial with both the first bevel gear and the epicyclic gear assembly for rotation about the first axis. Advantageously, the epicyclic gear arrangement is arranged between the first and the second bevel gear, wherein the epicyclic pinion gears mesh with both the first and the second bevel gear. As described, the main transmission has a continuously variable transmission, wherein the rotational speed of the output means varies according to the rotational speeds of the drive means according to the following formula: Vout = 2 × Vsecondary - Vprimary, where Vout is the output speed, Vsecondary the speed of the secondary drive (the epicyclic gear) and Vprimary the speed of the prime mover.

Preferably has the drive means of the differential gear assembly of the power transmission unit Differential pinion on which radially in the main gear to Arranged to be rotatable perpendicular to the first axis extending axes are. Advantageously, the output means of the differential gear assembly two differential cone side wheels on, coaxially with the differential pinions meshing about the first axis rotatable are arranged.

Preferably have the first and second bevel gear of the main gearbox respectively a centrally formed and axially extending hole. Preferably are the two differential side gears of the differential gear assembly centered on opposite ends of two coaxially aligned ones Power take-off parts, such as semi-axles, attached, extending through the holes extend axially outwardly in the first and second bevel gears of the main transmission and advantageously with effect the drive wheels are connectable.

Advantageously, the main transmission is operatively connected to the differential gear assembly by means of a differential frame which is connected to the second bevel gear of the main gear and carries the differential pinion of the differential gear assembly. As described, the differential gear assembly functions as a conventional automotive differential transmission. This embodiment is particularly advantageous for use with a primary and secondary engine and / or drive branches such as the real hollow shaft rotors or turbines in flat design or Hohlwellenprimär- and -sekundärantriebszweigen consisting of gears, sprockets or axial and radial pistons, Hydraulikrückhalte- and -energieversorgungseinrichtungen for controlling the reaction energy between the common engine and the primary and secondary drive branch of the singular or hybrid power source.

The Drive branches can expediently be arranged on each side of the transmission, the outputs through the respective primary and secondary drive branches Escape to an extremely simple, compact and lightweight To create a powertrain.

If the engine has petrol or diesel engines, is the Power shift transmission unit expediently centered in, below or on the same side as the engine and the primary and Secondary drive line, the above Hollow gears, sprockets or radial or axial hydraulics with hollow centers, as described, drive.

in the Case of a front-wheel drive is the power shift transmission unit expediently with respect to the engine and the drive branches between the front wheel assemblies, so that it with respect to the position and Room requirements with the present Vehicles from mass production is in line.

at a vehicle, a tractor or a lorry with four-wheel drive is the power transmission unit expediently with respect to the Row engine and the drive branches arranged so that the outputs of the powershift transmission through the respective primary and secondary drive branches and through the axially extending hole of the first and second Bevel gear of the main gear extend so that aligned with each other Power take-off parts such as half-axles in front of and behind the differential gear assembly extend, in turn, by means of universal devices and torque tubes forward and backward the front and rear differentials and Axes for all four wheels supplied with rotational energy.

at a two-wheel drive tractor or a truck with rear-wheel drive The engine can be moved in the normal, conventional way in the longitudinal direction stay organized, and the primary and Secondary drive lines supply this in the normal position of the standard differential located power shift directly with energy. Thus, at Heavy-duty high-torque trucks and tractors leading to optimization the speed and torque of the engine under heavy load and Torque requirements over a wide range of operating conditions usually 10- and 12-speed manual require that the manual transmission be omitted by the standard pinion and pinion just an extra Drive bevel gear and pinion is added, with the pinion itself within an annular Support of the outer gear parts can rotate freely. this makes possible a stepless speed range from reverse to overdrive by just the relationship between the primary and the secondary drive branch changed becomes.

If the engine a rotary engine, a real rotary engine, a gas turbine engine or having an electric motor, the power transmission unit expediently centrally between the engine on one side as primary drive branch drive and the secondary drive branch, as described, on the opposite Side, be attached, with the power take-off parts through the center of the hollow rotor or turbine shafts and through the hollow secondary drive system on the opposite side Extend side of the transmission unit. Furthermore, the housing of the power transmission unit one piece with the common housing the engines and the secondary drive branch be educated.

advantageously, For example, an internal combustion engine may have a direct primary drive branch and a secondary drive branch provide variable speed from a common shaft, forms the two drive branches with variable speed.

Of the Advantage of this arrangement is that a high-volume standard engine from mass production as the only source of energy or in some applications one Flywheel assembly as a means for storing kinetic Energy can be used. The flywheel can go to the engine belong or not. The flywheel assembly may be at times of maximum demand to complete of the power entry and / or for regenerative purposes Provide brakes. The flywheel can be directly from the single energy source powered or regenerative braking by means of a fixed Gear drive and a one-way clutch between the engine and the flywheel be used, the flywheel then over a common output shaft of the primary and the secondary drive branch energy supplies. Alternatively, the flywheel may be in a high speed position in FIG the primary or secondary drive branch with a suitable translation slow to the primary or secondary drive input be positioned of the transmission.

The The powershift transmission unit described above advantageously combines the functionality integrated with the main gearbox and the differential gear assembly such that the Drive rotation energy differential with continuously variable output speeds transmitted to two rotary power outputs will, while the single energy source with one speed and one torque is operated so that a maximum fuel efficiency, minimal pollution and a even performance of the vehicle results.

The Control device advantageously has a microprocessor control unit with an input device for receiving input commands, e.g. from the driver of the vehicle, and multiple input / output interface devices for one closed return scheme the performance parameter of the vehicle powertrain. The several Input / output interface devices advantageously have several high performance sensors for monitoring, analyzing and sending of data regarding the power parameter of the powertrain. Preferably, the from the microprocessor control unit continuously controlled performance parameters and performance parameters specific to the type of engine, including the oil pressures of the engine variable secondary Drive branch and the interactive forces and controls that the Drive train include. For example, if the engine is off a combustion piston engine, can the performance parameters continuously controlled by the microprocessor control unit Further, such performance parameters are specific to the most efficient one Fuel combustion and the least pollution are how For example, the pressure in the intake manifold and / or the boost pressure, the engine torque, the engine speed, the fuel-air mixture, the fuel flow, the sprocket timing, the valve setting, the variable Intake manifold geometry, variable compression, variable Pre-combustion chamber compression in case of indirect ignition Diesel engine, the combustion chamber conditions, the compression ratio and the exhaust gas chemistry and temperature.

in the The microprocessor control unit advantageously provides use for one shot self-diagnosis feedback loop control for continuous monitoring, Analysis and synergistic adjustment of performance parameters in Reaction to input commands of the driver. In particular, reacts the microprocessor control unit advantageously adapting to Input commands and / or analysis of the data regarding the performance parameters and continuously controls the speed and torque of the engines and the other engine variables to the fuel combustion efficiency maintain and keep pollution to a minimum, being the final one Output speed and the power of the drive train continuously be controlled to meet the operating requirements by synergistic Adjustment of performance parameters including relative speeds the power input means and the primary drive branch and the hydraulic pressures and speed of the variable secondary Drive branch and the division of the reactive load between the transmission and the two drive branches and the engine to meet.

advantageously, is the microprocessor control unit with a performance algorithm programmable, so that they the controlled performance parameters continuously according to the algorithm adapts to optimize powertrain performance. The microprocessor control unit For example, it may be programmed to optimize driveline efficiency in which case the microprocessor control unit is in response to a command input of the driver of the vehicle continuous monitoring and analysis as well as a synergistic adjustment of the performance parameters each of the two drive branches would perform from a single engine to maintain the efficiency of the engine within peak ranges, at the same time the load distribution between the engine and the primary Drive branch and the hydraulic pressure continuously monitored and is set and a reduction in the input speed of the secondary drive branch allows is in accordance with the transmitted from the transmission reactive load through change the length the trapped oil columns one Perform reduction, for the final Output speed and power of the transmission are synergistic to control them meet the operational requirements. In itself, the total benefits become of the engine over reached a wide range of different operating conditions.

It should be noted that significant improvements in fuel economy and correspondingly significant reductions in exhaust emissions are achieved when the engine is of the internal combustion engine type. It will also be appreciated that to maximize the engine's overall efficiency, the engine includes a rotary engine, a balanced and expansion cycle balanced rotary engine, a gas turbine engine, a diesel engine, a gasoline engine, an electric motor, or a hybrid engine and energy storage and regenerative braking system combination can. This could be done, for example, in the form of an internal combustion engine and electric generator with a lightweight high energy efficient battery storage and electric motor or internal combustion engine with combined brake energy recovery by a flywheel with a generator and a combined electric motor / generator flywheel for energy storage and regenerative braking as an engine. As another example, the invention may be used with a combination of hybrid engine technologies, such as in a fuel cell that converts methanol to hydrogen, where the hydrogen may be fed through a proton exchange membrane fuel cell in which it can be combined with atmospheric oxygen to produce energy for an electric motor to deliver, in turn, an electric motor / genes ratorschwungrad for energy storage and regenerative braking drives. This engine can provide power to the improved transmission as described, providing the means for interactive, real-time, accurate control of energy consumption and regeneration for optimum performance, eliminating fouling and expelling only water vapor. Alternatively, a fuel cell and regenerative braking could power efficient energy storage batteries to provide electrical power to an electric motor that is connected to the improved transmission in the described form either directly or through an energy storage flywheel. The transmission microprocessor control would immediately adapt the external, widely varying load requirements to an optimal drive power consumption matching the engine to prevent overloading the system but deliver optimal energy output.

at a further advantageous use of due to the reactive Power of the transmission in the secondary drive branch pressurized trapped oil can the energy storage takes place and the regenerative braking energy be stored by the oil Compressed nitrogen in a nitrogen-filled accumulator, the in short-term cases Maximum needs, as in stop / start situations in city traffic, energy supplies.

In an internal combustion engine burns the ignited fuel-air mixture in a flame front, extending to the extent of the limits of the cylinder expands. The resulting expanding gases press the Pistons away from the cylinder head, causing the crankshaft to rotate and a working stroke is generated. When the piston is expanding from the Gases moved away, increased he the cylinder volume. The higher the initial one Cylinder compression ratio is, the faster the burn rate and the corresponding Expansion rate and the stronger is the need for a higher piston speed, those of the higher ones Speed of the engine equals. For a motor with a fixed compression is it desirable in view of optimum cylinder pressure that the Piston speed to the rate at which the gases expand, is tuned, so that the best fuel efficiency and lowest pollution levels result. The intent of the preferred embodiments of this invention is to use the control of variable oil columns to change of the relationship the gear reaction forces to enable about the relationship between the prime mover branch and the secondary To vary drive branch of a common energy source. This will increase the reaction forces by the initial drive, by a common drive shaft is canceled, and the primary and secondary drive branch speed change, the to the outer gear is applied, continuously adjusted to the driving power source to enable for maximum Benefits the optimal balance between speed and torque to meet the variable load requirements and with the to maintain the desired Vehicle speed required energy levels in line to be. Because it, as described, has the function of a continuously variable transmission, wherein exactly the required vehicle speed results while the optimum fuel-air combustion rate is maintained, that the engine speed is varied and the compression ratios and air-fuel ratios can be varied over the full range of vehicle speed and energy requirements optimal operating conditions are maintained in real time, resulting in Minimal pollution and maximum fuel efficiency results. Another control of the speed of the flame front, especially with respect to a diesel engine, can be adjusted by an adjustable Pre-combustion chamber and / or reach the main combustion chamber. The interaction of the inputs and outputs from the microprocessor unit allows optimal settings the engine speed, the combustion chamber compression ratio and / or Vorverbrennungskammerhubraums, so that a quiet steady fuel-air combustion speed (Flame front) gives the mechanical movement and speed the retaining parts in a piston or rotary engine corresponds.

in the The following will now be embodiments of the invention, by way of example only, with reference to FIGS attached Drawings described.

SUMMARY THE DRAWINGS

1 FIG. 12 is a graph of the efficiency of an internal combustion engine relative to the variation in torque and engine revolutions per minute. FIG.

2 FIG. 12 is a gear shift diagram showing the performance pattern of an automatic transmission showing gradations in power input and vehicle speed. FIG.

3 Fig. 4 is a graphical representation of the comparison between a multi-stage conventional mechanical gear transmission and the powertrain transmission of this invention, wherein the reaction force controlled by the trapped oil in the form of a reverse hydro-transmission on the second the drive branch is used.

4 . 5 and 6 FIG. 12 is a graph of the relationships and reaction forces associated with the primary and secondary power output branches and the output of the external gearbox. FIG.

7 is a diagram and a partially sectioned schematic view of the illustrated ratios and reaction forces in linear form, which are converted in the outer gear in a circumferential shape.

8th Fig. 3 is a partially sectioned schematic view of one embodiment of a power transmission unit according to the first aspect of the invention covering the outer and inner gears and the ratio and effect of the primary and secondary power drive branches.

9 is a sectional view of a practical embodiment of the in 8th illustrated power transmission unit in relation to the primary power drive branch and the variable secondary power drive branch of a single engine.

10 FIG. 12 is a partially cut-away schematic view of one embodiment of the differential hydraulic / mechanical variable secondary power drive branch reactively connected to and controlling the outer power transmission shown in a maximum displacement position. FIG.

11 is a partially sectioned view of the in 10 shown secondary power drive branch in zero displacement position.

12 is a partially sectional view of the in the 10 and 11 shown secondary power drive branch, in which the full displacement range and the resulting control of the ratios and reaction forces is shown, which are directly connected to the primary and secondary power drive branches to the external powershift transmission, and the resulting outputs.

13 is a sectional view of a practical embodiment of the in 12 partially illustrated power shift transmission unit and the correlation of one in the 10 . 11 and 12 3, the differential output hydraulic / mechanical variable power secondary drive train and the primary power drive branch suitable for universal positioning and use in front wheel drive, rear wheel drive, and four wheel drive vehicles, tractors, and trucks, including direct differential and gearshift replacement, as far as cars are concerned , Trucks and tractors with front and rear axle drive is located in the conventional position of the differential.

14 is a partially sectioned schematic view of the Otto and diesel engine embodiment of the drive train according to the second aspect of the invention, wherein the correlation of the gasoline or diesel engine, the in 12 illustrated primary and secondary power drive branch and the in 13 illustrated power shift transmission in a preferred embodiment with power take-offs, which are located with respect to the engine at a position that is best suited for vehicles with front-wheel drive and cross-mounted engine, is shown.

15 is a partially sectioned schematic view of a simplified embodiment of the in 14 shown powertrain, which is suitable for the rear power take-off for conventional in-line engine positions in vehicles, trucks and tractors.

16 is a sectional view of a practical embodiment of the in 13 illustrated power transmission unit, which combines an embodiment of the differential displacement of the dynamic hydraulic / mechanical variable secondary power drive branch and the primary drive branch integral with the outer gear and the inner differential.

17 is a partially sectioned schematic view of an embodiment of an in 16 shown power shift transmission with an integrally formed primary and secondary drive branch.

18 Figure 12 is a graphical representation of the forward gear versus a selected gear in a prior art 24-speed transmission.

19 is a diagram of a drive train according to the prior art, which is used to achieve the in 18 shown 24 reduction ratios is required.

20 is a sectional view of a 24-speed transmission in 19 shown type.

21 is a sectional view of a torque converter in the transmission of 20 used type.

22 is a graphical representation of the Fluctuation in fuel consumption, torque and engine rpm performance in a typical diesel engine.

23 shows an embodiment of the drive train of this invention, which is suitable for high torque loads.

24 shows the application of an in 23 shown embodiment of the invention.

25 shows the functional part of the in the 23 and 24 shown arrangement.

26 shows a sectional view of a four-wheel drive tractor having a drive train according to an embodiment of the present invention.

27 shows a sectional view of a tandem drive assembly containing a drive train according to the present invention.

28 shows a further embodiment of the drive train according to this invention.

BEST TYPES THE EXECUTION THE INVENTION

First, the principle of operation of the part of the invention forming the transmission with reference to the 4 to 9 described.

DIFFERENTIAL RELATIONSHIPS AND REACTIONS

DRIVE RATIO 3 : 1

The representation of 4 indicates that during a forward movement of the rack 1 in the direction of arrow A by 30 teeth and with forward movement of the spindle 4 of the gear 2 in the direction of arrow B by a distance, the 10 teeth on the rack 1 equals what is equal to a power drive ratio of 3: 1, the gear 2 thereby moved by twenty teeth in a clockwise direction, causing a movement of the rack 3 in the direction of arrow C is effected by 20 teeth. However, the gear has 2 moved by the distance of 10 teeth in the direction of arrow B, the rack moves 3 The net effect is in a ratio of 3: 1 when a prime mover drives the rack 1 moved forward and a secondary drive the pivot 4 of the gear 2 at one-third the speed of the primary drive rack 1 moved forward. The rack 3 moves at one third of the input speed of the primary drive rack 1 in reverse direction.

The reaction forces associated with the 3: 1 ratio are that the rack 1 Force in the direction of arrow A requires and the spindle 4 of the gear 2 with a force in the reverse direction of the arrow B must be retained to cause the gear 4 around the axis 2 turns and a force to the rack 3 is applied, which then provides a force in the direction of arrow C. It can be seen that during a movement of the rack 1 and gear spindle 4 in the same direction with a ratio of 3: 1 the rack 3 usable work in the reverse direction, but only as long as the gear spindle 4 in the reverse direction to arrow B and in the same direction as the backward in the direction of the arrow C performed work is retained.

DRIVING RATIO 2 : 1

If the rack 1 in 5 by the distance of 30 teeth in the direction of arrow A is moved forward and if the gear spindle 4 around a 15 teeth of the rack 1 moving forward in the direction of arrow B, this results in a 2: 1 ratio between the primary and secondary power drives. If the gear spindle 4 only half the distance of the rack 1 has been found, it is found that the gear 2 clockwise about 15 teeth around the spindle 4 has turned, if the gear spindle 4 however, moving forward by a distance equal to 15 teeth, the rack remains 3 stationary. It can be seen that in the same direction with a ratio of 2: 1 moving primary drive rack 1 and secondary drive branch gear spindle 4 the rack 3 remains stationary in a dynamic neutral locking position.

The reaction forces with a 2: 1 ratio between the primary power drive branch rack 1 and the secondary power drive branch to the gear spindle 4 are that no other reaction forces are required than those required to overcome the friction, provided that the output rack 3 , which is stationary at C, does not act on any load.

When in the direction of arrow B, a load on the rack 3 is applied, a force-like same reaction load of the gear spindle B in the B opposite direction is required to compensate for the force applied to C, and on the rack 1 In the direction of the arrow A applied load must be equal to that on the gear 2 around the spindle 4 to be reverse rotating force acting from the rack 3 is transferred and equal to the force acting on this at C is. The net effect is that on the rack 1 in the direction of the arrow A the same force must be equal and that about the spindle 4 rotating gear 2 to the rack 3 is transferred so that it is equal to the rack 3 in the direction of arrow B is applied force, so that the spindle 4 the sum of the two on the rack 1 and the rack 3 is exposed to applied forces and in the direction opposite to the arrow B must be retained. That is, the spindle B requires twice the restraining force from the load at C on the driven rack 3 applied force.

If at C on the power take-off rack 3 in the direction opposite to the arrow B counter load is applied requires the spindle 4 the double force in the direction of arrow B and the rack 1 experiences the reaction force in the direction opposite to arrow A.

It can be seen that the rack 3 at a 2: 1 ratio of the primary and secondary power drive branches remains stationary, but depending on the direction of the on the rack 3 applied load even with rotating in the same direction primary and secondary drive branches the reaction energy of the applied load on the output rack 3 if reversed, an immediate reaction load on the primary drive branch rack 1 and the secondary drive branch gear spindle 4 in the opposite direction, so that the propelling of restraint loads is reversed and every time the load is shifted to the power take-off rack 3 at C changes.

DRIVE RATIO 1 : 1

In case the primary power arm drive rack 1 in 6 is advanced by a distance of 30 teeth in the direction of arrow A, and if the secondary power drive branch is the spindle 4 Moved in the direction of arrow B, it becomes clear that the sprocket 2 through the rack 1 , the gear spindle 4 and the gear 2 traveled equal distance and the rotation preventing engagement of the teeth of the gear 2 is retained in the teeth of the rack 1 from turning. That's why the rack is moving 1 and the sprocket 2 both as a solid mass at the same speed in the direction of arrows A and B. It is clear that when engaging the teeth of the gear 2 in the rack 3 This must also move at the same speed in the direction of arrow C.

For most vehicle functions with a 1: 1 ratio, the transmission is only occasionally subject to reverse thrust on the rack 3 suspended in the direction of arrow C ( 6 ). Then the vehicle slows down by using the compression of the engine as a brake, or by using an exhaust brake. At any other time is the rack 3 subjected to varying degrees of force in the direction opposite to the movement and arrow C. This force is generated by the engagement of the teeth of the gear 2 into the teeth of the rack 3 limited. The from the rack 3 on the gear 2 In the arrow C opposite direction applied force tries the gear 2 clockwise around the spindle 4zu Turn, what at a ratio of 1: 1 of the rack 1 is restricted, wherein a restraining force in the direction opposite to the arrow A ent is required. The on the spindle 4 force required in the direction of arrow B is the sum of the vehicle drive load on the rack 3 in the direction opposite to the rack-and-pinion motion, which is opposite to the arrow C, and the force required to drive the gear 2 by means of one on the rack 1 in the direction of movement and the direction opposite to the arrow A applied restraining force to rotate about the spindle 4 to prevent.

It can be seen that at a ratio of 1: 1, that at the secondary drive branch on the spindle 4 required force is in the direction of the arrow B and equal to twice the force of the performed on the power take-off rack C in the direction opposite to the arrow C work and that the force on the primary drive branch rack 1 is present in the same direction and equal to the force of acting against the power take-off rack 3 in the opposite direction to arrows C and A, is work.

RATIO 1: 1.5

As an overdrive, any ratio may be chosen over 1: 1, but for purposes of illustration a ratio of 1: 1.5 has been chosen. If the primary drive branch rack 1 by a distance of 30 teeth in the direction of arrow A is moved forward and the spindle 4 by a distance of 1.5 times that of the rack 1 moving forward, which corresponds to a total of 45 teeth, will, as in 6 shown the gear 2 When using this ratio, turn 15 teeth counterclockwise as the spindle 4 in the gear 2 has moved 15 teeth further than the rack 1 , The net effect is that the power take-off rack 3 travels the same distance of 30 teeth as the primary drive rack 1 plus the extra distance of 30 teeth caused by the rotation of the sprocket 2 is caused counterclockwise, with the rack 3 moved by a total distance of 60 teeth in the direction of arrow C and thus an overdrive of twice the speed of the primary drive branch is effected.

For illustrative purposes shows 7 the primary power drive branch as a linear rack 1 , Wherein the primary drive branch is denoted by A. The secondary power drive branch is as a spindle 4 and gear 2 represented, which is free around the spindle 4 can rotate and is referred to as a secondary drive branch B. The impact of the combination of ratios and reaction forces is related to 4 . 5 and 6 with the resulting power output in 7 at the rack 3 is denoted by C. The planar linearity of the racks and gears has been used for purposes of illustration, and it should be understood that the same ratios and reaction forces occur when the racks are used 1 and 3 by drive bevel gears 5 and 6 and the gear and the spindle 4 through a pinion 7 and a warehouse 8th are replaced.

If the rack 1 in the direction of arrow A is moved forward and the drive bevel gear 5 from the top, as seen in the same direction as the rack is moving, the drive bevel wheel rotates counterclockwise in the arrow direction on the drive bevel gear 5 , This causes the pinion 7 turning clockwise as seen from above, showing the same effect as the gear 2 with a flat profile; the pinion 7 turns the pinion gear 6 in the direction of the arrow on the power take off pinion gear at C.

As in the 4 . 5 and 6 explained, moves the secondary power drive branch according to 7 the spindle 4 and with it the gear 2 in the direction of the arrow at B. The same effect is achieved by turning the pinion 7 together with the bearing assembly 8th around the central axis of the drive bevel gears 5 and 6 ie around the axis AC.

It will be understood that although the ratios and reaction forces have been explained only in terms of the ratios 3: 1, 2: 1, 1: 1 and 1: 1.5, the output speed variation and reaction forces are proportional to the drive means, and though according to the following formula: Vout = 2 × Vsecondary - Vprimary, where Vout is the output speed, Vsecondary the secondary drive speed (the epicyclic gear arrangement), and Vprimary the speed of the prime mover, where the output force of the vector analysis corresponds to the reaction forces of the two drive means.

The invention allows an infinite stepless speed change between all ratios of reverse gear over stationary dynamic blockade over maximum speed and to overdrive. It is understood that the reaction forces generated between the designated with A primary power drive branch and the designated B secondary power drive branch must be limited so that they react against a common drive shaft and thus avoid energy loss. To accomplish this, the invention enables a speed variation between the power drive branch A and the drive branch B to be controlled by means of the differential change in the displacement of the columns of trapped oil or fluids in a rotational arrangement to give a differential hydraulic counteraction. whereby variable speed is provided and the reaction forces are controlled by the pinion 7 that is free in the camp 8th rotates, and the rotation of the pinion bearing assembly 8th about the axis of the drive bevel gear 5 and 6 around the axis AC according to 7 be generated.

This is in 8th to see more clearly. The primary power drive branch is with the bevel gear 5 , marked with arrow A, operatively connected. The secondary power drive branch is controlled by the differential variable speed output and the variable reaction torque by means of the dynamic differential hydraulic countermeasure of the trapped oil or fluid and provides for elimination or prevention depending on the reaction forces on the bearing 8th in which the pinion 7 can rotate freely as a counter reaction torque designated B, which rotates about the axis AA. Another aspect related to the invention and disclosed in Applicant's earlier applications is a differential 9 that in the outer gear 10 is included. The output speed and force are related to the speeds of the drive means according to the following formula: Vout = 2 × Vsecondary - Vprimary, wherein Vout is the output speed, Vsecondary the speed of the secondary drive (Umlaufgetriebeanordnung) and Vprimary the speed of the prime mover, the output force of the vector analysis of the reaction forces of the primary drive branch A and the secondary drive branch B corresponds, the output to the bevel gear 6 transferred and referred to as output power C. About a rotatable carrier 20 divide the bevel gear 6 the rotational force to the central differential and is denoted by D.

The invention has the advantage that proven pinion gears, pinions and differential components, which are in many different vehicles, load cars and tractors have been tried and tested, but used in a unique and novel combination with means for controlling speeds and reaction forces.

9 shows the gear unit 10 in accordance with the preferred embodiments of the first aspect of the invention, generally an outer main transmission 11 with two rotary drive means operatively connected to a rotary output means and an inner differential gear assembly 9 having rotational drive means operatively connected to two differentially rotatable output means. These components will now be described in more detail by explaining various preferred embodiments.

The main gearbox 11 the gear unit 10 advantageously has a gear train. The two drive means of the main gearbox 11 have a first bevel gear 5 and an epicyclic gear arrangement 12 coaxially arranged to rotate about a first axis AA. The epicyclic gear arrangement 12 has an annular pinion carrier 13 on, the inside arranged bevel pinion 7 rotatably supported with perpendicular to the first axis AA extending axes. As indicated by the letters A and B in 9 displayed are the first bevel gear 5 and the annular pinion carrier 13 each individually by means of conventional means such as a gear, a belt, a chain or directly operatively connected to the Leistungsantriebsprimär- and -secondary branch. The output of the main gearbox 11 has a second bevel gear 14 on, which is coaxial with both the first bevel gear 5 as well as with the epicyclic gear arrangement 12 is arranged for rotation about the first axis AA. The first and the second bevel gear 5 . 14 of the main gearbox 11 each have a centrally formed and axially extending hole. The epicyclic gear arrangement 12 is between the first and the second bevel gear 5 . 14 arranged, with the bevel pinions 7 both with the first and with the second bevel gear 5 . 14 comb.

The differential gear arrangement 9 is inside the main gearbox 11 arranged. The drive means of the differential gear arrangement 9 the power transmission unit 10 have differential bevel pinions 15 on the radially inside of the main gearbox 11 are arranged so that they rotate about perpendicular to the first axis AA extending axes. The output means of the differential gear assembly 9 have two differential cone side wheels 16 . 17 on, the coaxial and with the differential pinions 15 are arranged so that they rotate about the first axis AA. The two differential side gears 16 . 17 are centered at opposite ends of the two coaxially oriented power take-off parts 18 . 19 , such as half-axes, mounted through the holes in the first and second bevel gears 5 . 14 of the main gearbox 11 extend axially outwards. The half-axes 18 . 19 are operatively connectable to drive wheels (not shown) or, in the case of four-wheel drive, the power shift transmission unit may be mounted centrally, with the semi-axles 18 . 19 then operatively connected to the front and the rear differential. The main gearbox 11 is with the differential gear arrangement 9 by means of a differential frame 20 operatively connected, which with the second bevel gear 14 of the main gearbox 11 connected and the differential pinion 15 the differential gear arrangement 9 wearing.

10 FIG. 12 shows a practical embodiment of a transfer case apparatus according to this invention used to provide the power drive primary and secondary drive branches to the transmission from a single power source. FIG. This device provides for differential displacement and dynamic mechanical / hydraulic control of variable revolutions per minute and control of reaction forces between the prime mover and secondary thrusters. The reaction of the forces is controlled without energy loss by reaction against a common part in which a singular power drive either directly or by conventional means such as a gear, a V-belt, a multiple V-belt (with a common flat outer connecting belt), a toothed belt, a Roller or a special high-speed or low-noise chain or by means of a hydraulic device with a single power source is connected.

In 10 For example, the engine, such as a combustion gasoline or diesel engine, is illustratively and in a preferred embodiment operatively by means of a gear 20 connected, that the balanced arrangement 21 turns the gear 21 is on the case 22 fixed, which the primary drive branch via the rotatable housing to a gear 23 transfers, which is fixed to the housing. The housing is from bearings 24 on the fixed shaft 25 held by means of a mounting device 26 is prevented from turning. At the opposite end of the housing is further by means of a bearing 27 supported. This will generate energy directly via the prime mover branch from the engine via the gearwheel 20 through the housing 22 to the gear 23 at A transfer. The output speed is determined by the gear train between the engine or the engine and the arrangement 22 and directly changing the revolutions of the drive engine with respect to the gear 20 determined at A.

A second wave 28 , which are separately rotatable and independent of the shaft 25 is, is from camps 29 supported and directly with the gear 30 connected to the hydraulic / mechanical differential energy through the gear 30 provides, at B, the power-drive secondary branch is provided.

The fixed shaft 25 has a stationary circular piston group 31 on that by the spline groove on the fixed shaft 25 is prevented from turning. The camps 29 supported rotatable shaft 28 is by means of a spline groove on the rotatable piston group 32 attached so that they the piston group 32 effective with the shaft 28 and the secondary drive branch gear or sprocket 30 connects to provide B with a reaction force. On the rotatable housing 22 is a common slot and Kommutatorplattenanordnung 33 between the piston group 31 . 32 appropriate. The common slot and commutator plate assembly 33 is on the case 22 attached and rotatable with this. This ensures that the kidney shaped high and low pressure slot arrangement on either side of the perimeter of the assembly 33 always with the fixed pivot points 34 both the reaction displacement rotational group 31 , the swash plate control swash plate 35 and the fixed displacement swash plate 36 relative to the rotary housing 22 remains aligned.

The invention provides a differential hydraulic / mechanical speed variation and torque force response control between the fixed primary power drive branch A and the variable secondary power drive branch B. This is achieved by the primary two-pinion sprocket 20 on the drive A to the housing 22 mounted and rotatable and operative for the purpose of primary power output to A directly with primary drive branch chain or gear 23 is connected, wherein the torque force and the rotational speed directly by the rotational speed and the torque of the drive engine with respect to the gear or sprocket 20 controlled at A and directly to the toothed or sprocket wheel 23 be transmitted at A. That of the secondary power drive B as described in the 4 . 5 . 6 . 7 and 8th transmitted reaction forces react directly against the chain or gear 30 at B, as in 10 shown. The gear 30 gets off the shaft 28 and the shooting group 32 prevented from rotating, which in turn by the control of the trapped oil in the cylinder 37 is prevented from rotating, that of the piston 38 moved with the fixed angle swash plate 36 is in contact when the reaction forces the piston group 32 relative to the fixed angle swash plate 36 turn, the trapped oil column becomes 37 from the piston 38 shifted when this is to cylinder position 39 rotates. The amount of rotation allowed depends on the variable angle of the swash plate 35 when the displaced oil is released over the kidney-shaped slot assembly on either side of the circumference of the commutator assembly 33 can be transmitted in relation to the fulcrums 34 the swashplate remains fixed and aligned with the housing 22 rotates. It is clear that the trapped oil columns in the cylinder 37 can be freely transmitted through the kidney-shaped opening, which extends circumferentially by nearly 180 ° and a transmission of the oil through the commutator 33 to the cylinders 40 and 41 allows, as the rotary piston assembly 31 stationary and the swash plate 35 by the rotation of the housing 22 is driven. The displacement space in the cylinder 40 increases as the swashplate becomes cylinder position 41 moved, so that the control of the transfer of trapped oil from the cylinders 37 to the expanding cylinders between 31 . 41 is possible. It becomes clear that when the primary drive gear 20 the housing 22 turns and then the swash plate 35 over the festival swirling fulcrum 34 relative to the housing 22 drives and when the swashplate control angle on position 1 is set equal to the fixed angle of the swash plate 36 is the displacement of the trapped oil columns in the rotary piston assembly 32 with the expanding displacement chambers in the fixed rotary piston assembly 31 is identical, so that the reaction forces from the outer gear to B against the gear 30 can react by the shaft 28 relative to the prime mover sprocket 20 and the housing 22 turns backwards. That is, at full load and full power drive with respect to the gear 20 at A the output speed at the gear 30 Zero is and the reaction forces without energy loss due to the direct hydraulic / mechanical variable control, which mechanically in the rotating mechanical housing 22 has an interaction, are completely limited.

11 shows essentially the same arrangement as 10 but with around the axis 34 by means of an operating position 4 located adjusting device controlled swash plate 35 , Since the primary drive branch force at A is the gear 20 and the case 22 turns, she also presses the swash plate 35 with the same ratio of 1: 1 regardless of the angular position of the swash plate 35 , If the angle of the swash plate 35 on position 4 is set at an angle of 90 ° to the centerline of the shaft 25 is arranged, it becomes clear that no change in the displacement between the piston chambers 40 . 41 enters and any flow of trapped oil from the cylinders 37 . 39 is equal to zero. When the power-drive primary branch via the gear 20 at A over the case 22 and the gear 23 in the primary branch drive 1 of the outer mechanical transmission according to the 7 . 8th . 9 is transferred, which react in the 4 . 5 . 6 . 7 and 9 Negotiated reaction forces at B against the gear 30 , as in 11 shown. The gear 30 is through the wave 28 and the lobe group 32 limited, the piston 38 the trapped oil column in the cylinder 37 due to the zero displacement change in the fixed piston assembly 31 not through the commutator arrangement 33 can transfer. The net effect is that the whole arrangement 21 itself as a unit with the primary power drive and output sprockets 20 . 23 and the secondary reaction power control sprocket 30 at B, all of which rotate together at a ratio of 1: 1. At this point, with zero oil flow, 100% of the power is mechanically transferred.

12 is also with the 10 and 11 however, in addition to illustration, shows five operating positions of the swash plate 35 , In position 1 the power output of the output gear or sprocket occurs 23 over the primary drive branch to the sprocket 42 the external gear at A in direct correlation with that of the engine as described in the 10 and 11 provided speed and torque. When the control angle of the swash plate 35 in tax position 1 is located, the function occurs according to the explanation of 10 on, passing through the annular pinion carrier 13 transferred at B and to the chain or gear 43 applied reaction forces control the reaction forces, which are directly related to the change in the input and output loads and the direction of the forces as described in the 4 . 5 . 6 and 7 change, maintaining zero rotation.

When the control angle of the swash plate 35 in position 2 The shift allows the control of the reaction forces on the sprocket or sprocket 30 with the secondary power drive branch at B which is one-third that of the primary power drive branch at A. This produces an input to the main transmission at a ratio of 3: 1, with one-third of the input speed being input to a reverse output via the transmission of speed and energy through the bevel pinion 7 to the bevel gear 14 at C is transmitted.

12 shows that when on a position 3 set angle of the swash plate is a shift that allows the secondary power drive branch at B one revolution at half speed of the primary power drive branch A. This has a ratio of 2: 1 for the main gearbox 11 entail, as in the 4 . 5 . 6 and 7 described, with a dynamically blocked neutral position with zero output from the main gearbox 11 is produced. 12 shows that in position 4 located swash plate 35 which is at an angle of 90 ° to the shaft 25 The result is a drive with a ratio of 1: 1 between the primary power drive branch A and the secondary reaction power control drive branch B, as shown in FIG 11 is described. Thus, the 1: 1 ratio of the drives, which controls the interrelationship of the speeds and interactive torque forces in the main transmission, produces a 1: 1 full forward drive from the transmission via the second bevel gear 14 at C, the net effect being that the oil transfer flux is different from that in 12 shown piston rotation group to the cylinder 37 Is zero. There is no movement of the piston rotation group 32 with respect to the housing and all components in the hydraulic / mechanical differential displacement reaction power control unit 21 rotate together, except for the fixed wave 25 and the solid piston group 31 who always stay stationary.

The provision of the same speed in the ratio 1: 1 at the primary drive branch A and B and the control of the reaction energy by the power control drive branch at B has a zero movement of all components in the main transmission 11 result. Thus, for applications in vehicles, trucks and tractors, it can be seen that at zero speed there is only a 2: 1 change in the primary and secondary power branch speeds, with the secondary power branch B half the speed of the primary power drive branch with a smooth transition to a 1: 1 Ratio makes, so that a maximum efficiency for adaptation to variable loads can be done immediately and precisely.

If, as in 12 shown the control position of the swash plate 35 to position 5 can be achieved, a ratio of 1: 1.5 can be achieved, which has an overdrive output from the main transmission at C result, which is equal to 1.5 times the speed of the primary power drive branch, as previously in 6 described is.

An in 12 shown another aspect of the invention is that the rotary housing 22 is filled with hydraulic fluid to a point at which causes the centrifugal force during rotation, that the oil pressure in the direction of the outer circumference of the rotary housing 22 and build up the trapped oil mass. In one embodiment of the invention, this pressure acts as a filling pressure for filling the radially arranged cylinder 40 . 41 . 37 . 39 and the cylinder arranged radially between them, which are also in the outer periphery of the piston group 31 . 32 and in the high pressure hydraulic fluid area. The usual highs and low pressure shut-off valves are used and in certain applications, an integrated filling pump can be used.

A (not shown) further embodiment of the invention provides protruding fins and / or rotary seals for the housing 22 prior to the transport of hydraulic oil from and into the housing 22 for filtering and cooling.

In another embodiment consist of the lobe group and / or the commutator ceramic material, a high-speed, high-pressure, low-lubrication and highly efficient performance.

The adjustment swash plate 35 can be controlled by means of a suitable arm and thrust bearings or by means of appropriately performed pressure control or by radio frequency or infrared control to control the swash plate angle within the rotary housing 22 to enable.

13 shows a practical embodiment of the combined arrangement of 9 and 10 including the bearings, seals, and mounting arrangements required for a functioning assembly. These components may be varied in practice, and although one skilled in the art will be fully aware of their function, a description of the invention is not required.

In use, the main transmission described above 11 a continuously variable transmission, wherein the output speed (the speed of the second bevel gear 14 ) is proportional to the sum or difference of the two drive speeds, that of the integration of the arrangement 21 and provide for a variable primary and secondary speed and torque drive branch, designated as A and B (taking the speeds of the first bevel gear 5 and the epicyclic gear arrangement 12 Taxes). Thus it becomes clear that when the speeds of the first bevel gear 5 and the epicyclic gear arrangement 12 are controlled, the speed of the second bevel gear 14 can be varied progressively from a design maximum in a first direction of rotation above zero to a design maximum in a second opposite direction of rotation. As also described above, the differential gear assembly operates 9 in the manner of a conventional vehicle differential gearbox. Thus, the power transmission unit described above combines 10 expediently the functionality of the main transmission 11 and the differential gear assembly 9 einstckig such that drive rotational energy with continuously variable output speeds differentially to the two semi-axes 18 and 19 can be transferred.

The operation of the power transmission unit described above 10 will be further clarified with reference to the in 9 illustrated exemplary embodiment, in which the first bevel gear 5 and the second bevel gear 14 of the main gearbox 11 each thirty teeth and the bevel pinions 7 of the main gear each have ten teeth.

Below is the effect of the different drive ratios in the same direction of rotation about the axis AA for the in 9 illustrated exemplary embodiment described.

1: 1 DRIVE RATIO OF FORMING A HYDRAULIC / MECHANICAL RESPONSE POWER CONTROL UNIT 21 ONE-WAY CONNECTED PRIMARY AND SECONDARY POWER DRIVES A AND B

If the first bevel gear 5 according to 9 Turned one revolution forward are thirty teeth of the bevel gear 5 been turned forward. If at the same time the annular pinion carrier 13 Each revolution bevel pinion remains rotated one revolution in the same direction about the axis AA 7 stationary. Consequently, the non-rotating bevel pinions 7 whose teeth are in the second bevel gear 14 intervene that this and the differential frame 20 to turn forward one full turn around the axis AA. The of the differential frame 20 carried differential pinion 15 are only from differential side gears 16 and 17 limited, which is effective with half-axes 18 respectively. 19 are connected to allow a normal vehicle differential action between two (not shown) driving wheels.

This in 13 The result shown is that at one between the first bevel gear 5 and the annular pinion carrier 13 through the in 13 shown differential hydraulic / mechanical reaction power control unit 21 provided drive ratio of 1: 1, the speed ratio through the power transmission unit 10 and to the differential frame 20 1: 1.

DRIVE RATIO OF 2: 1 BY INTEGRATION WITH DIFFERENTIAL PERFORMANCE CONTROL UNIT 21 AND SPEED GEAR UNIT 10

If the first bevel gear 5 in 9 rotated one turn forward, moving forward by thirty teeth, and the annular pinion carrier 13 is rotated in the same direction by half a revolution about the axis AA (2: 1), has each bevel pinion gear 7 at fifteen tough ne about axes perpendicular to the axis AA extending axes rotated backwards and at half turn at the point of engagement on the second bevel gear 14 moved by fifteen teeth. Because the second bevel gear 14 has thirty teeth and the annular pinion carrier 13 and the bevel pinions 7 half a turn around, remain the thirtzahnige second bevel gear 14 and the differential frame 20 stationary.

As that of the power control unit 21 according to 13 between the first bevel gear 5 and the annular pinion carrier 13 with 2: 1 given drive ratio, remain as a result, the second bevel gear 14 and the differential side gears 16 and 17 in an effectively blocked stationary position with zero output by the power transmission unit 10 ,

DRIVE RATIO 3 : 1

If the first bevel gear 5 is rotated by thirty teeth forward by one turn and the annular pinion carrier 13 rotated one-third of a revolution in the same direction and the axis AA, rotates each bevel pinion 7 in the opposite direction by twenty teeth to the rear, taking an equivalent distance of ten teeth around the second bevel gear 14 return. Therefore, the second bevel gear rotate 14 and the differential frame 20 ten teeth or one third in the backward direction. As described above, the differential gear assembly allows 9 a normal vehicle differential action between two drive wheels (not shown).

With a drive ratio of 3: 1 between the first bevel gear 5 and the annular pinion carrier 13 As a result, the second bevel gear rotate 14 and the differential frame 20 by a third of a revolution in the reverse direction about the axis AA.

It it should be noted that Overdrive through a drive ratio of less than one. In terms of to the above description, it should be noted that the embodiments the integrated power transmission unit of the present invention a compact, one-piece combination a continuously variable transmission with a conventional one Provide vehicle differential, which is suitable for use in vehicle drive trains is, in particular, but not exclusively, in Vehicle drive trains, but also for vehicles with front-wheel drive, where they are in vehicles with four-wheel drive, trucks and tractors between the axles is mounted and wherein gearbox omitted and with the invention by rear differential drive assemblies in vehicles, trucks and tractors are replaced. In particular, the embodiments make In the transmission of the present invention, it is possible for two vehicle drive wheels to be differential with continuously variable speeds or front and rear axles be driven differentially.

As in the attached 8th . 9 and 13 In general, a vehicle powertrain according to the preferred embodiments of the second aspect of the invention generally includes integral integration of a singular power drive unit having a differential hydraulic / mechanical reaction power control unit 21 and the power transmission unit 10 with their integration and combination described above.

In 14 are a gear unit 10 with a torque responsive secondary hydraulic / mechanical power branch unit 21 with differential displacement and variable speed and a microprocessor control unit 45 shown. In the following, these components will be described in more detail, with various preferred embodiments will be explained.

In the following description, reference is made to the above detailed description of the construction, arrangement and operation of the various components of the power transmission unit 10 Referenced.

In the in 14 illustrated embodiments, the engine 44 from an internal combustion engine, alternatively any similar or dissimilar combination of any engine types from the conventional internal combustion engine sector - including the gasoline engine, the diesel engine, the turbo engine, the balanced rotary engine, the gas turbine engine or conventional types of electric motors, electric motors powered by fuel cells, electrical and / or mechanical Flywheels - can be used. Petrol and diesel internal combustion engines are preferred because they represent existing technologies at relatively low manufacturing costs. At the in 14 illustrated embodiment, the engine 44 a gasoline or diesel engine on.

Conveniently, the differential hydraulic / mechanical power control unit 21 arranged to drive via a matching gear, chain or belt from the crankshaft in the rear of the engine to a toothed or sprocket 20 transmits and the primary and secondary power drive branch to the center of the engine / engine 44 and via gears or chains or belts from the primary and secondary drive the power control unit 21 at A and B to the drive tooth or sprockets 42 and 43 the power transmission unit 10 transmits which with respect to the transverse engine 44 is centrally located at a position that is best suited for front-wheel drive applications, wherein the semi-axes 18 and 19 are in the proper position for connecting standard front wheel drive shafts to the front wheels.

The housing of the power control unit 21 and the powershift transmission 10 are also conveniently arranged so that they easily on the engine 44 can be attached or removed from this.

15 provides a simplified embodiment of the in the 9 and 10 illustrated vehicle drive train 46 in which the second bevel gear 14 the power transmission unit 10 is not operatively connected to an inner differential gear assembly, that is, the power transmission unit 10 only has the main gearbox 11 but instead is with a drive shaft 47 operatively connected. The simplified embodiment of the in 15 illustrated vehicle drive train 46 advantageously uses gasoline engines, diesel engines, turbo engines, balanced rotary engines or gas turbine engines and is designed so that it can be used in vehicles with front-wheel and / or rear-wheel drive, wherein the drive shaft 47 is operatively connected to a single conventional vehicle differential gear assembly. Alternatively, the in 15 illustrated vehicle drive train 47 be mounted in vehicles with four-wheel drive between the axles, the drive shaft 47 is operatively connected to front and rear conventional four-wheel drive differential gear assemblies.

In the 14 shown microprocessor control unit 45 has an input device for receiving a command input from the driver and a plurality of input / output interface devices (not shown) providing feedback loop control of the vehicle powertrain performance parameters. Preferably, the plurality of input / output interface devices (not shown) include a plurality of high performance sensors for monitoring, analyzing, and transmitting data regarding powertrain performance parameters. Advantageously, the performance parameters include the load on the engine 44 , the respective speeds of the two primary and Sekundärleistungsabtriebszahnräder or -kettenräder 42 . 43 , the load on the power control unit including the hydraulic pressure and the speed, the respective load on the two half-axles 18 and 19 and the respective speeds of the two half-axes 18 and 19 , Preferably, they include continuously from the microprocessor control unit 45 controlled performance parameters and performance parameters specific to the power unit (s) having the vehicle driveline.

In use, the microprocessor control unit sees 45 advantageously, a feedback loop controller for continuously monitoring, analyzing, and synergistically adjusting the performance parameters in response to the driver's command input. During the engine start 44 controls the microprocessor control unit 45 expediently the disengaging slip clutch and / or the angular orientation of in 12 shown swash plate, so that the secondary power drive branch is in a ratio of 2: 1 with respect to the primary drive branch, it being ensured that the driven bevel 14 of the main gearbox 11 at startup or a parked safety position remains in a stationary dynamically locked position.

That is, the microprocessor control unit 45 when starting only the respective power output shaft of the engine 44 with the first gear 20 the control unit 21 connects when the ratio of rotational speeds between the power take-off gear 42 and 43 2: 1 is.

From the above description of the operation of the powershift transmission units 21 and 10 shows that each of the differential side gears 16 and 17 the power transmission unit 10 is positively locked in this way in a secure stationary position at a speed of zero. It should be noted that in most applications to increase the safety level, a manually operated overrunning clutch is provided. From the above description of the operation of the powershift transmission units 21 and 10 also shows that the microprocessor control unit 45 in response to a subsequent command input, eg from the driver, the operation of the engine 44 and the power control unit 21 appropriately controls the relative speeds of the power take-off gears 23 and 30 adjust so that the differential side gears 5 and 14 the power transmission unit 10 and thus the two half-axes 18 and 19 and the two drive wheels (not shown) rotate in a desired direction at a desired speed. During operation of the vehicle powertrain 21 and 10 the microprocessor control unit responds 45 adaptive to a command input and / or analysis of the data related to the performance parameters and controls the final output speed and power of the vehicle driveline 48 continuously to meet operational requirements satisfy the performance parameters, including the relative speeds of the two primary and secondary power drive branches at A and B, and the reactive load control between the two drive branches 23 and 30 at A and B, be set synergistically.

Advantageously, the microprocessor control unit 45 programmable with a performance algorithm so that it continuously adjusts the controlled performance parameters synergistically according to the algorithm to match the performance of the vehicle powertrain 48 to optimize. The microprocessor control unit 45 for example, to optimize the efficiency of the vehicle powertrain 48 be programmed, the microprocessor control unit 45 In this case, in response to the command input, the respective performance parameters of the primary and secondary power drive branches would be continuously monitored, analyzed, and synergistically adjusted to increase the efficiency of the engine 44 while maintaining the load distribution between the primary and secondary drive branches of the power control unit 21 continuously monitored and adjusted to the final output speed and power of the power transmission unit 10 to synergistically manage to meet operational requirements. As such, overall engine benefits can be achieved over a wide range of different operating conditions. Thus, it should be noted that significant improvements in fuel economy and correspondingly significant reductions in emissions of exhaust gases can be achieved in embodiments employing internal combustion engines.

For example, if the engine is a combustion piston engine (as in the in 14 In addition, the performance parameters continuously controlled by the microprocessor control unit advantageously include parameters specific to the internal combustion engine, such as boost pressure, engine torque, engine speed, fuel mixture, spark timing, valve timing, variable intake manifold geometry, combustion chamber conditions, compression ratio, and exhaust chemistry.

Optimum operating conditions can be maintained in real time across the full range of speed and power requirements for vehicles, trucks and tractors, minimizing fouling and maximizing fuel efficiency by providing the function of a variable speed continuously variable transmission providing the exact vehicle speed required while maintaining the optimum fuel-air burn rate by maintaining the engine speed and the primary and secondary power output speeds of the power control unit 21 varies with respect to each other and by varying the fuel-air ratios.

In In view of the above description, it should be noted that embodiments of the vehicle powertrain of the present invention has a compact, provide adaptively controlled vehicle powertrain, in particular suitable for applications is where weight and size are critical are, as for example in small and medium-sized vehicles with front wheel drive.

With regard to the two drive means of the main transmission 11 and the differential gear assembly 9 for forming the power transmission unit 10 combined, embodied 16 the same principles as in 9 are described. The between the bevel gear 5 driven directly by the primary power drive branch and the annular pinion carrier 13 required speed change develops in the same way and controls the same reaction forces as in the 4 . 5 . 6 and 7 described.

16 shows a compact unified combination of the power control unit 21 the invention and the power transmission unit 10 which controls by differential displacement both the speed change and the above-described reaction forces. At the primary power drive sprocket or pulley 42 is a curved track 49 (a) attached, and the annular cam track is on a Kommutatorplatte 50 attached, the two kidney-shaped commutator slots 51 and 52 which extends nearly 180 ° radially on both sides of the circular commutator plate, with sufficient clearance between the opposed semicircular commutator slots to prevent movement of the hydraulic oil from the high pressure side to the low pressure side, said pressure being applied forces, in the 4 . 5 . 6 and 7 described, changed. In the 16 shown commutator slots 51 and 52 are each permanently fixed so that they with the cam plate 49 (a) . 49 (b) are aligned so that the highest part of the circular curved path 49 (b) at the dividing point of the two semicircular commutator slots 51 and 52 located. Because of this, regardless of the rotation of the primary drive chain or gear, the cam track rotates thereon 49 (a) . 49 (b) is attached, the commutator valve plate 50 and the commutator slots 51 and 52 also with the cam track to ensure a permanent alignment of the cam track and commutator plate for receiving and releasing hydraulic fluid, as the piston rotating group 53 This turns differentially when behind the piston 54 in the cy soothing 55 Hydraulic fluid or oil is included. The reaction force against the curved path 49 (a) . 49 (b) causing the piston rotating assembly to coincide with the primary power drive gear or sprocket 42 rotates. The piston rotating group is on the annular pinion carrier 13 attached, the inside arranged bevel pinion 7 rotatably supported, which cause an output speed change and torque forces, which in the manner previously described by the lobe group 53 be limited.

In order to control the two semicircular commutator slots, which are either a slot or a series of circular holes in a cam track which may also be connected to a narrow central slot, it is necessary to disconnect the high pressure commutator slot from the low pressure side and the commutator slots 51 and 52 also turn differentially, is a stationary Kommutatorgang Port A and a port on Port B. 57 necessary, which are offset relative to the axis AA against each other. This leaves the slots 51 and 52 always close to their respective high and low pressure ports, although the piston rotary cylinder oil passages 51 and 52 be rotated differentially over the commutator plate.

The advantages of the invention as described in 16 consist in that the differential control of the required change in the power drive speeds between the primary and secondary power drive branches and the control of the interactive forces is achieved in a compact integrated unit which replaces the differential at the current positions of all vehicles, trucks and tractors or any Such machines, which normally require a forward and reverse gear, variable speeds and differential action, is suitable and also allows the omission of the gearbox.

The complete integrated unit as described in 16 is a differential reactive hydraulic / mechanical engine 58 shown in this embodiment, wherein the reaction mechanism for controlling the forces in the secondary power drive branch is a piston equipped with a ball running on a circular concave cam track. It is intended to use other embodiments using pistons as well, which may be equipped with piston rings and rollers mounted at the end of the pistons by means of a bearing and a shaft about which the rollers are free to rotate and to one in Appropriately profiled cam track attack. This arrangement may also be equipped with side rollers to prevent the transmission of high torque forces to the pistons. An embodiment is later in 23 and another embodiment in 24 shown.

The invention as described in 16 comprises an external hydraulic controlled differential power control unit having a central main power transmission unit 11 and an internal drive differential 9 ,

It should also be noted that in case of application to vehicles, the differential carrier normally rotates in a ratio of approximately 4: 1 to the engine speed. For this reason, the invention of a unified differential reactive hydraulic / mechanical power shift transmission rotates 58 The long life and durability of all components half at full engine speed of about 4000 rev / min only 1000 rev / min.

By the lobe group 53 is placed on the outer circumference, it ensures high torque at low speed and also guarantees that the width between the semi-axes 18 . 19 for mounting between the front wheels in a front-wheel-drive vehicle to a minimum or without the semi-axles in normal space and normal axles as a replacement for the standard differential in vehicles, trucks and tractors, then a continuously variable transmission integral with a microprocessor control device in the described Form is provided.

In 16 have been described the speed and reaction forces, which culminate in that the control by the hydraulic fluid pressure and the flow rate at port A 56 and port B 57 is affected. In 17 is the primary power drive of the engine with a chain or gear 20 engaged, in turn, in a chain or gear 42 engages the gear 20 on a wave 62 attached to the camps 63 supported and in turn a Schrägachsenhydraulikverstellpumpe 69 which, in turn, outputs and receives hydraulic fluid via Port A 60 and Port B 61.

The opening 64 the sloping axis hydraulic variable pump goes from zero to maximum flow via the rotation of the control screw 65 controlled, in turn, continuously from the engine 66 controlled by a command from the in 14 shown microprocessor control unit is moved. If the ports 60 . 61 the Schrägachsenaxialkolbenhydraulikverstellpumpe in a suitable manner with the ports 67 . 68 in the combined differential reactive hydraulic / mechanical power shift transmission 58 The control of the reaction forces and speed of the piston rotation group can be connected 53 performed in the manner described above who the.

The invention as described in 17 can also with the Schrägachsenaxialkolbenverstellpumpeneinheit 69 which are operatively associated with a similar arrangement to a commutator plate 50 paved curved path 49 connected and mounted so that it rotates about a shaft on the axis AA, wherein the piston group 53 is capable of being differentially relative to the cam plate 49 and their supporting parts, which also rotate about the axis AA, the unit 58 via the offset ports, as described above, with the bent axis pump assembly 69 is connected, but both units than from the power transmission unit 10 separate unit in the same way as in 13 described operatively with the power transmission unit 10 are connected. Another embodiment of the invention is that the rotary piston group is in use in the manner described above in the form of axial or radial pistons acting against cam plates, so as to give the effect described, and wherein axes provided with flat sides, with Axles provided with curved sides or balls or rollers with or without axes (such an embodiment is later in 23 . 24 shown) against the proper closed cam track in either an axial or a radial position.

In various embodiments, the invention is suitable for all vehicle, truck and tractor types. For tractors and trucks, reference is made to the requirements and application of the invention to tractors and trucks, both of which provide exceptional torque requirements and are extremely exposed to external, highly fluctuating energy requirements. Examination of the currently available and most advanced technology and methods for dealing with the high loads and extreme variations in torque requirements is useful for better understanding the operation and application of the invention in tractors, trucks, and other types of high performance transportation machines. 18 shows a graphical representation of a 24-speed transmission, with the at 69 1, which is a forward gear at a speed of 3.4 km / h, and which passes through gradations that allow the engine to go through repeated cycles at each gear, with each cycle running at a few revolutions per minute (rpm) and a second high torque load request begins, and for the engine at low speed, a point is reached at which generates little power and consequently in the conversion of the fuel-air mixture in power high fuel consumption and low efficiency. As the speed of the engine increases, the power output increases until the engine speed and torque are balanced. At this point, which is generally in the middle of each gear, fuel economy is most efficient. As the engine speed increases in a particular gear, fuel consumption increases and efficiency in converting fuel to power decreases. In 18 This cycle represented by the graph repeats itself through 24 gears and engine cycles in the form described above. The curve 70 sets the desired average gear ratio as variable, allowing the engine to constantly balance between engine speed and torque load, with fuel consumption lowest at that point and engine efficiency highest in converting fuel to power. The object of the invention is to make it possible for the engine at any time to run with maximum efficiency in this compensating state.

19 is a diagram of how to achieve in 18 24-speed shifts required complex powertrain.

The driveline diagram in 19 starts at the engine 71 , The torque and horsepower (power) of the engine are through a torque converter 72 To assist in balancing the shock loads generated during the gear change and the manual or automatic engagement of the clutches, the torque converter is also used in particular at the start, which is associated with low speed and heavy load. The efficiency aspect of torque energy transfer is another cause of power loss due to the dissipation of energy dissipated as heat. In a tractor according to 19 apart from the 24 forward gears, this is a reversible power-shift set 73 , a drive coupling 74 , six drives 75 and a creeper / operation 76 required.

20 FIG. 12 is a cross-sectional view of a 24-speed transmission showing the complexity and high cost of a 24-speed transmission or ratios for matching the power curve to an acceptable level. The invention has for its object to achieve a stepless change in the gear ratios using a minimum number of wearing parts and lower investment costs, while increasing the efficiency of the powertrain, reduces the shock loads and excludes engine stalling or overshoot and improves fuel savings and the Pollution can be reduced.

20 shows the turbo coupling / torque converter 76 , the power shift reversal set 77 , the drive coupling 78 , six drives 79 , Creepers / operations 80 , the four-wheel drive output to the front axle 81 ,

21 is a sectional view of the turbo coupling / torque converter 76 ,

22 is a graphical representation of the effect of the balance between speed and torque, in terms of fuel efficiency, on a motor. The baseline 89 shows the actual speed range of a diesel engine when used on a tractor. The torque level is in Newton meters 83 and a graph of torque is included 84 shown. The performance is at 85 expressed in kilowatts and by a power curve at 86 shown. The effects of engine speed and torque on fuel economy are included in the graph 87 to be seen as a consumption curve. It should be noted that at a low speed of 1000 rpm at 88 and a high torque of 640 Newton meters of the engine fuel combustion is 215 g / kWh. Engine speed is between 1400 rpm and 1600 rpm 89 and the torque at 650 Newton meters, fuel consumption is at its lowest point of 195 g / kWh 92 , When the engine speed at 2200 rpm at 90 increases and the torque increases 85 Newton meter at 91 is reduced, the fuel consumption increases 92 to 225 g / kWh too.

It it becomes clear that the The object of the invention is the engine at any time, regardless of strong fluctuating loads and power demands, in the peak efficiency range to operate, which is the point of balance between speed and torque for this particular engine is, as in the graph at 1450 rpm and a torque of 640 Newton meters with a fuel combustion of 195 g / kWh is shown.

The graphic representation in 22 is used for illustrative purposes only, as the invention provides for compensation of engine speed and torque with widely varying speed and torque ranges, optimally selected and continuously reactive through inputs and control of the outputs of the microprocessor control unit according to US Pat 14 is changed in the previously described form.

23 is an embodiment of the invention, which is suitable for high-torque heavy loads in vehicles and in particular in heavy transport by trucks and high torque operating loads in forward and reverse for tractors.

Although the main gear 11 dimensioned according to the high load and torque requirements, it embodies in terms of the two drive means of the main gearbox 11 and the differential gear assembly 9 for forming the power transmission unit 10 combined, the same, in 9 described principles. The bevel gear driven mechanically between the primary power drive branch 5 and the annular pinion carrier 13 required speed change develops and controls the same reaction forces as those in the 4 . 5 . 6 and 7 . described

23 shows a compact unified combination of the hydraulic / mechanical differential power control unit as part of the invention in combination with the power shift transmission 10 that, similar to the description of 16 , the speed change and the reaction forces in the manner previously explained controls.

23 explains, in particular, the compact nature and high-torque properties of this embodiment of the invention. The primary power drive branch 42 is on an annular support member 92 attached, which in turn forms a support and on an annular, internally wound cam 93 is attached. In 25 is shown a side view of a portion of the cam track and axial piston assembly. In 23 is the cam 93 on an apertured valve housing 94 attached, which is together with the cam 93 rotates and thus ensures that regardless of the rotational position, the hydraulic fluid passages 95 and 96 are correctly aligned and for the supply of high pressure oil and the return of oil and for the correct position with respect to the in 25 shown cam holes 97 to care.

The stationary housing 98 in 23 is at the non-rotating part of the valve body 99 appropriate. The ports 100 and 101 are connected to the corresponding separate internal gears to group with the rotary lobe 102 to be in communication so that high pressure and return oil in the cylinder chambers 103 (a) and 103 (b) is transferred, at the appropriate time to the piston 104 (a) and 104 (b) drive out and on the rollers 105 (a) and 105 (b) attack so as to be engaged at the correct position on the curved path to the linear thrust of the piston 104 (a) . 104 (b) convert into rotational movement and / or as a retaining unit for controlling the rotational speed of the annular pinion carrier 13 with respect to the bevel gear 5 to act. This is achieved by linear and radial separation of the gears 95 and 96 achieved, including the linear gears, with the respective cylinder chambers 103 (a) . 103 (b) and the other cylinders which radially and uniformly in the rotatable piston group are ordered, of which a section in 25 shown is related.

The effect of the rotary lobe group 102 directly on the annular pinion carrier 13 is attached, which is the internally arranged recirculating bevel pinions 7 rotatably, thereby causing an output speed change and torque forces, which are limited by the fact that the rotary lobe group 102 Hydraulically differentially against the curved path in the manner previously described 93 acts.

In 23 the hydraulic differential speed control takes place by means of axial piston rollers 105 (a) . 105 (b) located on the outer circumference of the power shift transmission driveline assembly 106 against the curved path 93 act to achieve maximum efficiency and maximum torque application.

The dynamic hydraulic differential power control is located on one side of the annular pinion carrier 13 to form a compact integrated and unified arrangement with the power transmission unit 10 ,

23 shows the invention in a compact and efficient Hochmomentform, which is suitable for all applications described above, but especially as a direct replacement of vehicle, truck and tractor differentials, to act both as an infinitely variable continuously variable transmission from reverse to overdrive and also the two axes 18 and 19 To supply differential energy. In an embodiment (not shown), the differential may be 9 a self-locking differential or a controllable differential lock order be. The invention in the in 23 described form can be advantageously positioned centrally in a four-wheel drive tractor, wherein the semi-axes 18 and 19 with universal devices and propshafts to the front and rear differentials and axle arrangements of a four-wheel drive vehicle or tractor according to 26 are connected.

24 is an embodiment of the in 23 described invention, wherein the mechanical primary drive branch through the pinion 107 and the bevel gear 108 to the bevel gear 105 and also about the gear and the sprocket 42 to the toothed or sprocket wheel 20 is transmitted, which in turn on the Schrägachsenverstellpumpe 69 attached, in turn, over the ports 60 and 61 that work with the ports appropriately 60 and 61 connected in a suitable way with the ports 100 and 101 in the combined differential reactive hydraulic / mechanical power shift transmission 106 as described in 23 are connected, supplying and receiving hydraulic fluid. The invention as described in 24 is particularly suitable as a replacement for transmission and differential and is in the normal position for differentials in vehicles and in particular trucks and tractors, being self-contained and requiring only one high-speed crankshaft stub for driving on A with the pinion 107 is connected, and wherein the other external input control by the input from the microprocessor control unit 45 as described in 14 on the motor-activated guide spindle 66 as described in 17 is.

As previously described, all are subject to differential hydraulic / mechanical Power control units, which in all figures with respect to these Invention, the reversal of the reaction forces and if necessary, with check valves equipped to introduce hydraulic fluid on the low pressure side allow the hydraulic circuit and so prevent cavitation. When the energy reversal occurs, can the shut-off valves on both sides of the circuit integral with pressure relief valves and / or a filling pump be the low pressure hydraulic fluid around the alternating low pressure sides of the circuit to be supplied in the required case.

25 works as before in 23 described. In 25 is the segment of the curved path 93 to see from the end, with the lobe group 102 , the cylinder chambers 103 (a) . 103 (b) and 103 (c) , the pistons 104 (a) . 104 (b) and 104 (c) , the roles 105 (a) . 105 (b) and 105 (c) you can see. The curved path 93 has a trajectory in the form of opposite waves, which is symmetrical and thus remains in balance. The pistons 104 (a) . 104 (b) . 104 (c) and the other pistons are radially arranged and activated so that they remain in balance when turning. The radial arrangement of the cylinders and pistons in the radial piston rotating group 102 are relative to the waveform of the curved path 93 arranged so that, while the pistons 104 (c) and the role 105 (c) in a position to transfer energy either to the waveform of the curved path 93 gone or gone, is the role 105 (b) and the piston 104 (b) have completed the cycle and the cylinder chamber 103 (b) Maximal displacement has experienced. This action takes place uniformly in stages over the multiple pistons and undulations around the entire circumference of the assembly to achieve a smooth transition from circulating to reciprocating oil columns progressing from the in-line 24 shown Schrägachsenverstellaxialkolbenpumpe 69 be transferred and controlled.

26 shows a sectional view of a four-wheel drive tractor with an embodiment 106 of the invention as described in FIG 17 , which is advantageously mounted so that it differential energy to the front and rear axle 108 and 109 supplies.

27 shows a sectional view of a tandem drive assembly 110 wherein the main transmission / manual transmission and the primary drive differential by an embodiment 111 the invention as described in the 23 and 24 are replaced.

The mechanical PTO shaft from the in-line engine provided power and would be on 109 connected, with the mechanical drive power of the motor in the direction of the arrow 112 took place. When the gear assembly is installed in the differential case as indicated by the arrow 113 is displayed, the transmission differential full axles would be at 114 and 115 connected and supplied both axes in the appropriate position 116 and 117 with delivered power. It should be noted that the integrated engine 111 in tandem and / or single-axle machines as described in 27 can be used.

28 shows an embodiment of the invention, wherein in the outer main transmission 11 as described in 9 the differential gear assembly 9 on one side of the outer main gearbox 11 with the annular pinion carriers 13 may be attached, which is the outside arranged bevel pinion 7 rotatably store. As a result, a power shift transmission unit 10 be provided with very high torque, wherein the reduced outer diameter of a maximum diameter of the main gear 11 allows that the first bevel gear 5 and the coaxially arranged circulation arrangement 12 for rotation about a first axis AA. The epicyclic gear arrangement 12 is arranged inside and on an inner annular pinion carrier 13 supported. From the hollow annular pinion carrier 13 stand spindles 4 from and are attached to this. The spindles 4 support the bevel pinions 7 who are using camps 8th free around the spindles 4 can turn. The outer ends of the spindles 4 are on a toothed or bevel gear 43 appropriate. The function of the gear unit 10 is in 8th . 9 described, wherein the primary drive branch in 28 as A, the variable secondary drive branch at 43 as B and the power output via the second bevel gear 6 is shown as C, which in turn is directly at D on the differential carrier 20 is attached.

This allows the output force via the second bevel gear 6 of the main gearbox 11 to one side of the main gearbox 11 to a differential gear assembly 9 can be transmitted with a large diameter, which is attached to one side and whose diameter is equal to that of the bevel gears 5 . 6 can be to allow a normal differential transmission of power at very high torque.

The normal differential function provides power over the side gear 17 to the axis 19 and over the side gear 16 through the hollow bevel gear 6 , the carrier assembly 118 and the bevel gear 5 by means of the axis 18 ,

It should be noted that the differential can be arranged left or right and the principle in the various embodiments, as shown in the 9 . 12 . 13 . 14 . 15 . 16 . 17 . 23 . 24 . 26 shown and described, can be introduced. In the 28 described invention is particularly for use in the in the 24 and 27 described embodiments, in particular as a direct replacement for a standard differential, whereby both the function of a gearbox and a differential in a limited radial space is made possible, which for the purpose of ground clearance, the reduction of the diameter is possible when use in applications with extremely high torque, such as For example, in tractors, trucks and high-performance vehicles according to 27 he follows. It should be noted that the use of the same technology as used in 28 As a unified improved powertrain, which combines the previously described embodiment of a hydraulically controlled differential hydraulic / mechanical control unit with a main power transmission unit and a laterally mounted differential to enable a continuously variable transmission for all types of vehicles, trucks and tractors and a differential function, only one question of scale.

Though the present invention has been particularly preferred embodiments described, but she lets herself with different changes and embody modifications, which may differ in particular from those who have been described in the above description. These Modifications and changes are without departing from the scope of the invention limited by the claims possible.

Claims (23)

Transfer gear device ( 21 ) for providing two rotary drives ( 23 . 30 ; 5 . 13 ) with a variable relative speed from a single rotary drive ( 20 ; 42 ), the apparatus comprising: a first rotation element ( 22 ; 42 ) which is driven in rotation about a rotation axis and a first rotation output ( 23 ; 5 ) having; a second rotation element ( 28 ; 13 ) which is rotatable about the rotation axis and a second rotation output ( 30 ; 13 ) having; a first fluid chamber ( 31 ; 56 . 57 ; 64 ) connected to the first rotation element ( 22 ; 42 ) belongs; a first controller ( 35 ; 51 . 52 ; 69 ) for varying the volume of the first chamber ( 31 ; 56 . 57 ; 64 ) in response to the rotation of the first rotary element ( 22 ; 42 ); a second fluid chamber ( 32 ; 55 ; 103a . 103b ) connected to the second rotary element ( 28 ; 13 ) belongs; a second controller ( 36 ; 54 ; 104a . 104b ) for varying the volume of the second chamber ( 32 ; 55 ; 103a . 103b ) in response to the rotation of the second rotary element ( 28 . 13 ); characterized in that a commutator ( 33 ; 50 ; 94 ) in response to the rotation of the first rotary element ( 22 ; 42 ) is rotatable to at least regularly closed fluid communication between the first and the second chamber ( 31 ; 56 . 57 ; 64 : 32 ; 55 ; 103a . 103b ) during the rotation of the first and second rotation elements ( 22 ; 42 : 28 ; 13 ) to produce; wherein the rotational speed of the second rotational element ( 28 ; 13 ) in response to the rotation of the first rotary element ( 22 ; 42 ) by corresponding changes in the volume of the first and second chambers ( 31 ; 56 . 57 ; 64 : 32 ; 55 ; 103a . 103b ) is determined. Contraption ( 21 ) according to claim 1, wherein either the first or the second controller ( 35 . 36 ) is selectively adjustable to the rotational speed of the second rotational element ( 28 ). Contraption ( 21 ) according to claim 1, wherein at least one of the first and second regulators ( 35 . 36 ) a piston ( 38 ), which in the first or second fluid chamber ( 31 . 32 ) is movable to vary the volume. Contraption ( 21 ) according to claim 1, further comprising: a third fluid chamber ( 31 ) connected to the first rotation element ( 22 ) belongs; a third controller ( 35 ) for varying the volume of the third chamber ( 31 ) in response to the rotation of the first rotary element ( 22 ); a fourth fluid chamber ( 32 ) connected to the second rotary element ( 28 ) belongs; a fourth controller ( 36 ) for varying the volume of the fourth chamber ( 32 ) in response to the rotation of the second rotary member; and wherein the commutator ( 33 ) selectively closes fluid communication between the first and third chamber pairs ( 31 ) and second and fourth chambers ( 32 ). Contraption ( 21 ) according to claim 4, wherein the first and third chambers ( 31 ) oppose each other substantially diametrically across the axis of rotation and the second and fourth chambers ( 32 ) oppose each other substantially diametrically across the axis of rotation. Contraption ( 21 ) according to claim 3, wherein the pistons ( 38 ) are movable in a direction which is substantially parallel to the axis of rotation, and the device further comprises a first and second swash plate ( 35 . 36 ) for the operation of the pistons ( 38 ). Contraption ( 21 ) according to claim 6, wherein at least one of the swash plates ( 35 . 36 ) is selectively adjustable to change the volume of a corresponding chamber of the first or second chamber ( 31 . 32 ) to control. Contraption ( 21 ) according to claim 1, wherein the commutator ( 33 ) comprises one or more openings surrounding the axis of rotation substantially on its circumference. Contraption ( 21 ) according to claim 8, wherein the openings are kidney-shaped. Contraption ( 21 ) according to claim 1, wherein the first rotation element ( 22 ) a hollow cylindrical housing ( 22 ) comprising the second rotation element ( 28 ) at least partially surrounds. Contraption ( 21 ) according to claim 10, wherein the first fluid chamber ( 31 ), the commutator ( 33 ) and the second fluid chamber ( 32 ) in the housing ( 22 ) are included. Contraption ( 21 ) according to claim 11, wherein the first fluid chamber ( 31 ) is fixed against rotation, the commutator ( 33 ) with the housing ( 22 ) and the second fluid chamber ( 32 ) with the second rotation element ( 28 ) turns. Contraption ( 21 ) according to claim 1, wherein the corresponding changes in the volume of the first and second fluid chambers ( 31 . 32 ) can be adjusted so that a common rotation of the first and second rotation element ( 22 . 28 ) provided. Contraption ( 21 ) according to claim 1, wherein the corresponding changes in the volume of the first and second fluid chambers ( 31 . 32 ) can be adjusted so that the second rotation element ( 28 ) at half the speed of the first rotary element ( 22 ) is rotated. Contraption ( 21 ) according to claim 1, wherein the fluid chambers ( 31 . 32 ) are filled with oil. Powershift transmission unit ( 10 ), comprising: a transfer case device ( 21 ), an outer main transmission ( 11 ) and an inner differential gear assembly ( 9 ); the main gearbox ( 11 ) two rotary drive means ( 5 . 12 ), both of which each from a first rotational output ( 23 ) and a second rotational output ( 30 ) of the transfer case device ( 21 ), wherein the two drive means ( 5 . 12 ) operable with the rotatable output means ( 14 ) are connected so that the speed of the output means ( 14 ) in relation to the algebraic mean of the rotational speed of the two drive means ( 5 . 12 ) may vary; wherein the differential gear assembly ( 9 ) in the main transmission ( 11 ) is arranged and a rotatable drive means ( 15 ) operable with two differentially rotatable output means ( 16 . 17 ), wherein the output means ( 14 ) of the main transmission ( 11 ) and the drive means ( 15 ) of the differential gear assembly ( 9 ) are operatively connected; characterized in that the transfer gear device ( 21 ) according to claim 1. Motor vehicle drive train ( 46 ) that is capable of being continuously controlled over a predetermined operating range, and comprising: a single engine ( 44 ); a power shift transmission unit ( 10 ), which is a transfer case ( 21 ), an outer main transmission ( 11 ) and an inner differential gear assembly ( 9 ); the main gearbox ( 11 ) two rotatable drive means ( 5 . 12 ), both of which each from a first rotational output ( 23 ) and a second rotational output ( 30 ) of the transfer case device ( 21 ), wherein the two drive means ( 5 . 12 ) operable with the rotatable output means ( 14 ) are connected so that the speed of the output means ( 14 ) in relation to the algebraic mean of the rotational speed of the two drive means ( 5 . 12 ) may vary; wherein the differential gear assembly ( 9 ) in the main transmission ( 11 ) is arranged and a rotatable drive means ( 15 ) operable with two differentially rotatable output means ( 16 . 17 ), wherein the output means ( 14 ) of the main transmission ( 11 ) and the drive means ( 15 ) of the differential gear assembly ( 9 ) are operatively connected; characterized in that the transfer gear device ( 21 ) according to claim 1. Powershift transmission unit ( 10 ) according to claim 16, wherein the first fluid chamber ( 31 ) is fixed against rotation, the commutator ( 33 ) with the housing ( 22 ) and the second fluid chamber ( 32 ) with the second rotation element ( 28 ) turns. Motor vehicle drive train ( 46 ) according to claim 17, wherein the first fluid chamber ( 31 ) is fixed against rotation, the commutator ( 33 ) with the housing ( 22 ) and the second fluid chamber ( 32 ) with the second rotation element ( 28 ) turns. Contraption ( 21 ) according to claim 1, wherein both the first and the second fluid chamber ( 31 . 32 ) a piston ( 38 ) in the first and second fluid chambers ( 31 . 32 ) is movable to vary the volume. Contraption ( 21 ) according to one of claims 1 to 3, further comprising a Hydraulikverstellpumpe ( 69 ) obtained by rotation of the first rotary element ( 42 ) is driven to the first fluid chamber ( 56 . 57 ) Supply pressurized fluid. Contraption ( 21 ) according to claim 21, wherein the hydraulic variable displacement pump is a bent axis pump ( 69 ). Contraption ( 21 ) according to claim 3, wherein the piston ( 54 ) through a curved track surface ( 49 ) on the first rotation element ( 42 ) is operated.